Refrigeration cycle apparatus

ABSTRACT

A refrigeration cycle is switched from a refrigeration cycle circuit to a hot gas heater circuit to make the hot gas discharged from a compressor flow into an evaporator and heat the vehicle passenger compartment, at which time, when a suction pressure (Ps) of the compressor becomes a low pressure below a first predetermined pressure, the discharge volume (Vc) of the compressor is made larger to ensure a sufficient auxiliary heating performance. Further, when the hot gas is made to flow into the evaporator to heat the vehicle passenger compartment, if the suction pressure (Ps). of the compressor becomes a higher pressure over a second predetermined pressure, the discharge volume (Vc) of the compressor is made smaller to protect the refrigeration cycle parts and lighten the ON, OFF shock.

This is a division of Application Ser. No. 09/126,802, filed Jul. 31,1998, now U.S. Pat. No. 6,148,632.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a refrigeration cycle apparatus forheating the interior of a vehicular passenger compartment, moreparticularly relates to a vehicular use air-conditioning system providedwith a refrigeration cycle apparatus designed to guide the hightemperature, high pressure gas phase refrigerant discharged from arefrigerant compressor into a refrigerant evaporator and heat the airflowing through a duct at that refrigerant evaporator.

2. Description of the Related Art

In the past, the general vehicular use heating system used has been ahot water type heating system which guides the coolant water which hadbeen used for cooling the engine into a heater core in a duct to heatthe air flowing through the duct by that heater core and thereby heatthe interior of the passenger compartment. This hot water type heatingsystem, however, suffered from the problem of a remarkably insufficientheating capacity when just starting up the engine and activating the hotwater heating system, that is, when the hot water heating system is juststarting up, when the temperature of the outside air was low and thetemperature of the used cooling water was consequently low.

To solve the above problem, for example, Japanese Unexamined PatentPublication (Kokai) No. 5-223357 has proposed a vehicular use airconditioning system (related art) provided with a refrigeration cycleapparatus (auxiliary heating system) designed to augment the heatingcapacity of the heater core by leading the high temperature, highpressure gas phase refrigerant (hot gas) discharged from the compressorof the refrigeration cycle apparatus through a pressure reducingapparatus to a refrigerant evaporator and heat the air flowing throughthe duct at that refrigerant evaporator. Note that the compressor was anengine-driven compressor driven by the engine through an electromagneticclutch.

During the heating operation, when the temperature of the cooling wateris over a predetermined temperature, the heating capacity of the heatercore of the hot water type heating system is sufficiently high, so thecompressor is turned off and the auxiliary heating system isdeactivated. Further, when the temperature of the cooling water is lowerthan a predetermined temperature, the heating capacity of the heatercore of the hot water type heating system is insufficient, so thecompressor is turned on to activate the auxiliary heating system.

Further, when the discharge pressure from the compressor is higher thana predetermined pressure, the load on the compressor is too high, so thecompressor is turned off to deactivate the auxiliary heating system andprotect the refrigeration cycle apparatus. Further, when the dischargepressure from the compressor is less than a predetermined pressure, thecompressor is turned on to activate the auxiliary heating system.

Further, in the apparatus of this related art, an accumulator isprovided between the refrigerant evaporator outlet and the compressorsuction side to separate the gas and liquid phases of the refrigerantand lead the gas phase refrigerant for use elsewhere so as to preventthe suction of liquid phase refrigerant to the compressor and thereforeavoid an adverse effect on the life of the compressor by liquidcompression.

In this refrigeration cycle apparatus of the related art, however, ifthe heating operation at the hot gas heater circuit-is continued for apredetermined time (for example, about 30 minutes), both the high-sidepressure and the low-side pressure of the refrigeration cycle apparatusbecame higher than during the cooling operation using the ordinaryrefrigeration cycle circuit. For example, the high-side pressure of therefrigeration cycle apparatus would become 20 to 25 kg/cm² during aheating operation (operation by a hot gas heater circuit) and 13 to 15kg/cm² at a cooling operation (operation by a refrigeration cyclecircuit). Further, the low-side pressure of the refrigeration cycleapparatus would become 4 to 5 kg/cm² during a heating operation and 1 to2 kg/cm² during a cooling operation.

Further, during a heating operation by the hot gas heater circuit,compared with the cooling operation at the normal refrigerationcycle-circuit, as mentioned above, both the high-side pressure and thelow-side pressure of the refrigeration cycle apparatus would becomehigher and the torque fluctuation would become greater when turning thecompressor from the on state to the off state. Therefore, when thecompressor was turned from the on state to the off state duringoperation of the vehicle, the rotational speed of the engine beltdriving the compressor would fluctuate tremendously and therefore theproblem would arise of deterioration in both the power performance anddriveability of the vehicle.

It may therefore be considered to control the capacity and control thepressure without frequently turning the compressor on and off bychanging the compressor to a cooler use variable volume type compressorsuch as used in the past. This conventional cooler use variable volumetype compressor, however, is designed to reduce the discharge volumefrom the compressor the lower the suction pressure to the compressor.

When installing such a cooler use variable volume type compressor in ahot gas heater circuit for a heating operation, the larger the heatingload, that is, the lower the temperature of the air sucked into theevaporator, the lower the temperature and the pressure of therefrigerant used for heat exchange with the air in the evaporator. Dueto this, since the discharge volume from the compressor becomes smallerdue to the variable volume control of the compressor, the flow of hightemperature refrigerant into the evaporator also becomes smaller andtherefore the problem arises that the auxiliary heating performance,that is, the performance in assisting the heating capacity of the heatercore, is no longer sufficiently manifested.

Further, when installing a cooler use variable volume type compressorinto the hot gas heater circuit for a heating operation, when theheating capacity is small, that is, when the temperature of the airsucked into the evaporator is high, the temperature and the pressure ofthe refrigerant used for the heat exchange with the air in theevaporator become high. Due to this, since the discharge volume from thecompressor becomes larger due to the variable volume control of thecompressor, the discharge pressure from the compressor becomes larger.

If the high-side pressure of the refrigeration cycle apparatus were torise to an abnormally high pressure temperature (for example, 27kg/cm²), the components of the cycle such as the refrigerant pipingwould malfunction or break. Further, even if the heating load is small,as explained above, if the high-side pressure of the refrigeration cycleapparatus reaches 25 kg/cm², the flow of the high pressure refrigerantinto the evaporator will also become large, so the problem will arise ofan excessive auxiliary heating capacity for assisting the heatingcapacity of the heater core.

Further, in the above apparatus of the related art, while no referencewas made to the specific configuration of the apparatus, as is generallyknown, a calibrated orifice passage for recovering the fine oil isprovided near the bottom of the inside of the accumulator. The liquidrefrigerant (including the lubrication oil) near the bottom of theinside of the accumulator is designed to be drawn in from thiscalibrated orifice passage so as to eliminate the insufficientlubrication of the compressor and protect the compressor life.

The present inventors engaged in actual experiments and studies on howfar the opening of the calibrated orifice passage (the passage diameter,also called the bleed port diameter) should be set and found that thefollowing problem arises. That is, when the opening degree of thecalibrated orifice passage is enlarged to the maximum value (forexample, ø2.5) at the time of heating operation in the winter (operationby hot gas heater circuit), the amount of suction of liquid refrigerantinto the compressor through the calibrated orifice passage increases, sothe amount of compression work of the compressor increases and it ispossible to increase the heating capacity as well. On the other hand,since the amount of suction of the liquid refrigerant at the time of acooling operation in the summer also increases, the amount of oilcirculating in the cycle also increases which invites a reduction in thecooling capacity and an increase in the power consumption of thecompressor.

Therefore, if the diameter of the calibrated orifice passage is reducedto the optimal value for a cooling operation in the summer (for example,ø1.2), the amount of suction of liquid refrigerant during a coolingoperation in the summer will fall and the cooling capacity will beimproved so it would be possible to reduce the power consumption of thecompressor, but at the time of a heating operation in the winter, theamount of compression work of the compressor will fall so the heatingcapacity will fall and become insufficient.

SUMMARY OF THE INVENTION

An object of the present invention is to provide a refrigeration cycleapparatus provided with a variable discharge volume means enablingachievement of a sufficient heating performance. Further, an object isto provide a refrigeration cycle apparatus able to prevent breakdown andbreakage of the refrigerant piping and other cycle components and ableto prevent an excessive heating capacity. Further, an object is toprovide a vehicular use air-conditioning system able to obtain anoptimal venting temperature by the minimum necessary power.

A further object of the present invention is to secure the capacity andreduce the power consumption of the compressor during a coolingoperation and, simultaneously, secure the capacity at the time of aheating operation.

According to a first aspect of the present invention, there is provideda refrigeration cycle apparatus comprising:

(a) a refrigerant compressor driven in rotation by an internalcombustion engine so as to compress the refrigerant,

(b) a refrigerant evaporator for performing heat exchange with air onthe inflowing refrigerant to cause it to evaporate and vaporize,

(c) a refrigerant circulation circuit for circulating the refrigerantdischarged by the refrigerant compressor to the refrigerant evaporatorand returning it to the refrigerant compressor, and

(d) variable discharge volume means for increasing the discharge fromthe refrigeration cycle apparatus when a suction pressure into therefrigeration cycle apparatus becomes lower than a predetermined value.

According to a second aspect of the present invention, there is provideda vehicular air-conditioning system comprising:

(a) an air-conditioning duct for leading air-conditioned air into avehicle passenger compartment,

(b) a refrigerant evaporator arranged in the air-conditioning duct forperforming heat exchange with air on the inflowing refrigerant to causeit to evaporate and vaporize,

(c) a heater core arranged in the air-conditioning duct at a downstreamside of air of the refrigerant evaporator for heating air using usedcooling water of an internal combustion engine as a source of heat forheating,

(d) a refrigerant compressor driven in rotation by an internalcombustion engine so as to compress the refrigerant,

(e) a refrigerant circulation circuit for circulating the refrigerantdischarged by the refrigerant compressor to the refrigerant evaporatorand returning it to the refrigerant compressor,

(f) a cooling water circulation circuit for circulating the coolingwater flowing out from the internal combustion engine to the heater coreand returning it to the internal combustion engine,

(g) a variable discharge volume means for reducing the discharge volumefrom the refrigeration cycle apparatus when the discharge pressure fromthe refrigeration cycle apparatus becomes higher than a setting,

(h) a venting temperature detecting means for detecting a temperature ofair vented from the air-conditioning duct into the vehicle passengercompartment, and

(i) a variable discharge pressure means for setting the dischargepressure lower the closer the venting temperature detected by theventing temperature detecting means is to a target value.

BRIEF DESCRIPTION OF THE DRAWINGS

The present invention may be more fully understood from the descriptionof the preferred embodiments of the invention set forth below togetherwith the accompanying drawings, in which:

FIG. 1 is a view of the overall configuration of a vehicular useair-conditioning system according to a first embodiment of the presentinvention;

FIG. 2 is a sectional view of an electromagnetic clutch and a variabledischarge volume type compressor according to the first embodiment ofthe present invention;

FIG. 3 is an explanatory view showing the general structure of anelectromagnetic volume control valve according to the-first embodimentof the present invention;

FIG. 4 is a block diagram of a control system of a vehicular useair-conditioning system according to the first embodiment of the presentinvention;

FIG. 5 is a flow chart of the method of control of the discharge volumeby an air-conditioner ECU according to the first embodiment of thepresent invention;

FIG. 6 is an explanatory view showing the state of operation of theelectromagnetic volume control valve at the time of a cooler modeaccording to the first embodiment of the present invention;

FIG. 7 is an explanatory view showing the state of operation of theelectromagnetic volume control valve at the time of a heater modeaccording to the first embodiment of the present invention;

FIG. 8 is an explanatory view showing the general structure of anelectromagnetic type volume control valve according to a secondembodiment of the present invention;

FIG. 9 is an explanatory view showing the general structure of anelectromagnetic type volume control valve according to a thirdembodiment of the present invention;

FIG. 10 is an explanatory view showing the general structure of anelectromagnetic type volume control valve according to a fourthembodiment of the present invention;

FIG. 11 is a flow chart of the method of control of the discharge volumeby an air-conditioner ECU according to the fourth embodiment of thepresent invention;

FIG. 12 is an explanatory view showing the general structure of anelectromagnetic type volume control valve, switching control valve, andhot gas volume control valve according to a fifth embodiment of thepresent invention;

FIG. 13A is a graph of the relationship between the suction pressure anddischarge volume of a compressor, while FIG. 13B is a graph of therelationship between the discharge pressure and discharge volume of thecompressor;

FIG. 14 is a view of the overall structure of a vehicular useair-conditioning system according to a sixth embodiment of the presentinvention;

FIG. 15 is an explanatory view of the general structure of anelectromagnetic type volume control valve, switching control valve, andhigh pressure control valve according to a sixth embodiment of thepresent invention;

FIG. 16 is a block diagram of the control system of a vehicular useair-conditioning system;

FIG. 17A is a graph of the relationship between the setting of thedischarge pressure of a compressor and the control current, while FIG.17B is a graph of the relationship between the discharge pressure anddischarge volume of the compressor according to the sixth embodiment ofthe present invention;

FIG. 18 is a view of a refrigeration cycle apparatus showing a seventhembodiment of the present invention;

FIG. 19 is an explanatory view of the valve mechanism in the seventhembodiment of the present invention;

FIG. 20 is a block diagram of the electrical control in the seventhembodiment of the present invention;

FIG. 21 is a graph of the effect of the seventh embodiment;

FIG. 22 is an explanatory view of the valve mechanism in an eighthembodiment of the present invention;

FIG. 23 is an explanatory view of the valve mechanism in a ninthembodiment of the present invention;

FIG. 24 is a view of the operating characteristic of the valve mechanismin the ninth embodiment of the present invention;

FIG. 25 is a longitudinal sectional view of an accumulator showing a10th embodiment of the present invention;

FIG. 26 is a view of a refrigeration cycle apparatus showing an 11thembodiment of the present invention;

FIG. 27 is a longitudinal sectional view of an accumulator showing a12th embodiment of the present invention;

FIG. 28 is a time chart showing the suction pressure and dischargepressure of a refrigerant compressor according to a 13th embodiment ofthe present invention;

FIG. 29 is a view of the configuration of a refrigeration cycleapparatus of a vehicular air-conditioning system according to a 14thembodiment of the present invention;

FIG. 30 is a sectional view of a variable throttling valve according tothe 14th embodiment of the present invention;

FIG. 31 is a graph of the opening degree of the variable throttlingvalve with respect to the high-side pressure of the refrigeration cycleapparatus according to the 14th embodiment of the present invention; and

FIG. 32 is a sectional view of a differential pressure valve accordingto a 15th embodiment of the present invention.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

FIG. 1 to FIG. 7 show a first embodiment of the present invention. FIG.1 shows the overall structure of a vehicular use air-conditioningsystem.

The vehicular use air-conditioning system of this embodiment is avehicular use air-conditioning system where the air conditioning means(actuators) in the air conditioning units 1 for air-conditioning theinterior passenger compartment of a vehicle carrying an engine (internalcombustion engine) as a main heat source for heating are controlled byan air-conditioning control apparatus (hereinafter referred to as anair-conditioning ECU) 10.

The air-conditioning unit 1 is provided with an air-conditioning duct 2constituting an air-conditioning passage 11 for leading air-conditionedair into the passenger compartment. At the upstream most side of the airin the air-conditioning duct 2 are provided an outside air suction port,an inside air suction port, and an inside and outside air switching door(none shown). At the further downstream side of the air is provided acentrifugal blower fan 3. Further, at the downstream most side of theair in the air-conditioning duct 2 are provided a defroster vent, a facevent, or a foot vent or other vent and mode door (not shown).

The centrifugal blower fan 3 is comprised of a scroll casing providedintegrally with the air-conditioning duct 2, a blower fan motor 12controlled by a not shown blower fan driving circuit, and a centrifugalblower fan 13 driven to rotate by the blower fan motor 12. Note that theflow rate of the centrifugal blower fan 13 of this embodiment isdesigned to be switched continuously or in stages from a 0 stage (OFF)to 32 stage.

Next, at the upstream side of the air from the vent, there is provided aheater core 5 of a hot water type heating system (main heating system)for reheating the air passing through a later mentioned evaporator 6.The heater core 5 is provided in the middle of the cooling watercirculation circuit 14 where a circulatory flow of cooling water isproduced by a water pump (not shown) driven by the engine E. The heatercore 5 is a downstream side heat exchanger (second heating use heatexchanger) through which the cooling water which has absorbed theexhaust heat of the engine E is recirculated when a hot water valve 15provided in the cooling water circulation circuit 14 and which uses thecooling water as a heat source for heating so as to reheat the air, thatis, performs an air heating action. The engine E, heater core 5, coolingwater circulation circuit 14, and hot water valve 15 comprise a hotwater type heating system 4.

Next, between the centrifugal blower fan 3 and the heater core 5 isarranged an evaporator 6 forming a component of the refrigeration cycleapparatus 20 mounted in the engine so as to block the entire area of theair passage 11 in the air-conditioning duct 2. The above refrigerationcycle apparatus 20 is provided with a first refrigerant circulationcircuit (hereinafter referred to as a refrigeration cycle circuit) 21, asecond refrigerant circulation circuit (hereinafter referred to as a hotgas heater circuit) 22, and first and second solenoid valves 23 and 24for switching between the refrigeration cycle circuit 21 and the hot gasheater circuit 22.

The refrigeration cycle circuit 21 is a refrigeration cycle apparatus inwhich the high temperature, high pressure gas phase refrigerantdischarged from the compressor 7 is recirculated from the first solenoidvalve 23 to the condenser (refrigerant condenser) 25, receiver(gas-liquid separator) 26, expansion valve (first pressure reducingmeans) 27, evaporator 6, accumulator (gas-liquid separator) 28, andcompressor 7 in that order. Further, the hot gas heater circuit 22 is arefrigerant circuit in which the high temperature, high pressure gasphase refrigerant (hot gas) discharged from the compressor 7 isrecirculated from the second solenoid valve 24 to the pressure reducingapparatus (second pressure reducing means) 29, evaporator 6, accumulator28, and compressor 7 in that order.

The refrigeration cycle apparatus 20 recirculates the refrigerant intothe refrigeration cycle circuit 21 when the first solenoid valve 23opens and the second solenoid valve 24 closes. Further, therefrigeration cycle apparatus 20 recirculates the refrigerant to the hotgas heater circuit 22 when the first solenoid valve 23 closes and thesecond solenoid valve opens. Note that the circulation circuit switchingmeans of the present invention is comprised by the first and secondsolenoid valves 23 and 24. Further, reference numeral 16 is a coolingblower fan which is driven by a drive motor 17 to forcibly blow outsideair to the condenser 25.

The evaporator 6 corresponds to the refrigerant evaporator of thepresent invention and functions as a cooling use heat exchanger whichevaporates the low temperature gas and liquid phase refrigerant flowingin from the expansion valve 27 to cool the passing air when therefrigerant flows in the refrigeration cycle circuit 21. Further, theevaporator 6 functions as a first heating use heat exchanger (hot gasheater of auxiliary heating system and auxiliary heat source system)which passes the high temperature gas phase refrigerant flowing in fromthe pressure reducing apparatus 29 to heat the air passing through itwhen the refrigerant flows through the hot gas heater circuit 22. Here,the expansion valve 27 not only insulates and expands the refrigerant,but also adjusts the amount of the refrigerant circulated in accordancewith the superheating of the refrigerant at the outlet of the evaporator6.

Next, the compressor 7 of the present embodiment will be brieflyexplained based on FIG. 1 to FIG. 5. Here, FIG. 2 is a view of avariable discharge volume type compressor formed integrally with anelectromagnetic clutch. The compressor 7 has connected to it anelectromagnetic clutch 8 which transmits or shuts off the power of theengine E to the compressor 7.

The electromagnetic clutch 8 is comprised of a stator housing 32 affixedto a housing 44 of the compressor 7 through an annular mounting flange31, a rotor 34 to the outer periphery of which is connected a pulley 33connected to the engine E by a belt V, an armature 35 arranged facingthe rotor 3 across a close distance and formed with a friction surfacefor frictionally engaging with the friction surface of the rotor 34, anelectromagnetic coil 37 for attracting the armature 35 to the rotor 34against the elastic force of the rubber hub (elastic body) 36 by thegeneration of a magnetic flux when energized, and an inner hub 39connecting the armature 35 and the shaft 40 of the compressor 7 throughan outer hub 38 and rubber hub 36.

The compressor 7 corresponds to the refrigerant compressor of thepresent invention. It is a known for example waffle type which canchange the discharge volume and is comprised of a shaft 40 rotatingintegrally with the inner hub 39 of the electromagnetic clutch 8, aswash plate 41 affixed to the shaft 40 at an angle, a piston 42 set onthe swash plate 41, a housing (front housing) 44 connected to a cylinder(rear housing) 43 through which the piston 42 slides, and anelectromagnetic volume control valve (corresponding to the variabledischarge volume means of the present invention) 9 connected to the rearend of the housing 44 and able to change the discharge volume of thecompressor 7.

Here, the cylinder 43 forms a cylinder chamber 45 with the piston 42.Toward the center of the baffle plate 46 forming the cylinder chamber 45is formed a suction port (not shown) opened and closed by a suctionvalve (not shown) formed by an elastic metal sheet. The suction port iscommunicated with a suction port 48 formed in the valve body 47 of theelectromagnetic volume control valve 9. Further, toward the outside ofthe valve plate 46 is formed a discharge port 50 opened and closed by adischarge valve 49 formed by an elastic metal sheet. The discharge port50 is communicated with a discharge port 51 formed in the valve body 47.Note that inside the housing 44 is provided a crank chamber 52 formoving the swash plate 41 to freely displace and fixed calibratedorifices 53 a and 53 b (see FIG. 3) for effectively communicating thesuction port 48 with the discharge port 51.

From the above, it is seen that when the electromagnetic coil 37 of theelectromagnetic clutch 8 is in the energized state (ON), the armature 35of the electromagnetic clutch 8 is drawn to the rotor 34 and the rotor34 and armature 35 frictionally engage, whereby the drive power of theengine E is transmitted through the belt V and the electromagneticclutch 8 to the shaft 40 of the compressor 7. By this, the refrigerationcycle apparatus 20 is activated and thereby the air cooling action orair heating action of the evaporator 6 is performed. Further, when theelectromagnetic coil 37 of the electromagnetic clutch 8 is deenergized(OFF), the armature 35 of the electromagnetic clutch 8 moves away fromthe rotor 34 and the frictional engagement of the rotor 34 and armature35 is broken. Due to this, the drive power of the engine E is nottransmitted to the shaft 40 of the compressor 7 and the air coolingaction or air heating action by the evaporator 6 is stopped.

Next, an explanation will be made of the electromagnetic volume controlvalve 9 based on FIG. 1 to FIG. 3. Here, FIG. 3 is a view of the generalstructure of an electromagnetic volume control valve 9 housed in thecompressor 7.

A refrigerant pressure circuit is formed in the body of the compressor 7and the valve body 47 of the electromagnetic volume control valve 9. Therefrigerant pressure circuit is comprised of pressure passages 54 to 56through which the suction pressure (Ps) of the compressor 7 is guided,pressure passages 57 and 58 through which the discharge pressure (Pd) ofthe compressor 7 is guided, a pressure passage 59 giving crank chamberpressure (Pc) to the crank chamber 52 of the compressor 7, a passageport 61 communicating with the communication passage 60, and acommunication passage 62 communicating with the pressure passage 59.Note that the communication passage 60 communicates the convergencepoint of the downstream side of the pressure passage 55 and thedownstream side of the pressure passage 58 with the communication port61. Further, the communication passage 62 communicates the convergencepoint of the downstream side of the pressure passage 56 and thedownstream side of the pressure passage 57 with the pressure passage 59.

The opening degree of the communication port 61 is determined by thestopping position of the valve element 63. The stopping position of thevalve element 63 is designed so as to be determined by the displacementposition of the plunger 64 and the bellows 65. The plunger 64 andbellows 65 are communicated with the valve element 63 through the rods67 and 68. The setting position of the plunger 64 is designed so as tobe changed by the magnitude of the control current to theelectromagnetic coil 69. Note that reference numeral 70 indicates areturn spring for returning the plunger 64 to its initial position.

The opening and closing of the pressure passages 57 and 58 aredetermined by the stopping position of the valve element 71. Further,the opening and closing of the pressure passages 55 and 56 aredetermined by the stopping position of the valve element 72 linked withthe valve element 71. The stopping positions of these valve elements 71and 72 are designed to be changed by the magnitude of the controlcurrent to the electromagnetic coil 73. Note that reference numeral 74is a return spring for returning the valve elements 71 and 72 to theinitial positions.

Therefore, the electromagnetic volume control valve 9 is a variabledischarge volume means for changing the discharge volume of thecompressor 7 by changing the setting of the suction pressure (Ps) of thecompressor 7 by the control current from the air-conditioner ECU 10.That is, the electromagnetic volume control valve 9 is constructed to beable to change the external force acting on the plunger 64 and thebellows 65 by applying the control current to the electromagnetic coil69 in the valve body 47. By changing the relationship of the openingdegree of the valve element 63 to the suction pressure (Ps), the actualpost-evaporator temperature (TE) is controlled to the targetpost-evaporator temperature (TEO).

Next, an explanation will be given of the air-conditioner ECU 10 basedon FIG. 1 and FIG. 4. Here, FIG. 4 is a view of the control system ofthe vehicular use air-conditioning system.

The air-conditioner ECU (heating control means) 10 for controlling theair-conditioning means in the air-conditioning unit 1 receives as inputthe switch signals from the switches on the air-conditioner operationpanel (not shown) provided on the front panel of the passengercompartment. Note that on the air-conditioner operation panel areprovided a mode selection switch 100 for switching the air-conditioningmode to either of a cooler mode (cooling operation) and heater mode(heating operation), a temperature setting switch (temperature settingmeans) 101 for setting the temperature in the passenger compartment to adesired temperature, an air-conditioner switch 102 for turning on or offthe refrigeration cycle apparatus 20, a blower fan switch 103 forturning on or off the centrifugal blower fan 3, etc.

Further, in the inside of the air-conditioner ECU 10 is provided a knownmicroprocessor comprised of a CPU, ROM, RAM, etc. Sensor signals fromsensors are converted from an analog to digital format by a not showninput circuit, then input to the microprocessor. Note that theair-conditioner ECU 10 is designed so that when the ignition switch (keyswitch) controlling the start and stopping of the engine E of thevehicle is turned on (IG ON) and DC power is supplied from the powersource in the vehicle, that is, the battery (not shown), the controlprocessing is started.

The air-conditioner ECU 10 receives as its input sensor signals from theinside air temperature sensor (inside air temperature detecting means)104 for detecting the air temperature in the passenger compartment(hereinafter referred to as the inside air temperature), an outside airtemperature sensor (outside air temperature detecting means) 105 fordetecting the air temperature of the outside of the passengercompartment (hereinafter referred to as the outside air temperature), asunlight sensor (sunlight detecting means) 106 for detecting the amountof sunlight entering the passenger compartment, a post-evaporatortemperature sensor (post-evaporator temperature detecting means) 107 fordetecting the air temperature directly after passing through theevaporator 6 (hereinafter referred to as the post-evaporatortemperature), a cooling water temperature sensor (cooling watertemperature detecting means) 108 for detecting the temperature of thecooling water flowing into the heater core 5, and a refrigerant pressuresensor (high-side pressure detecting means) 109 for detecting thehigh-side pressure (discharge pressure: Pd) of the refrigeration cycleapparatus 20. Note that the above switches and sensors detect theair-conditioning environmental factors required for air-conditioning thepassenger compartment of the vehicle.

Next, a brief explanation will be made of the control of the compressorcapacity by the air-conditioner ECU 10 of the present embodiment basedon FIG. 1 to FIG. 5. Here, FIG. 5 is a flow chart showing the method forcontrol of the discharge volume by the air-conditioner ECU 10.

When the ignition switch is turned on (IG ON) and DC power is suppliedto the air-conditioner ECU 10, the routine of FIG. 5 is started. First,the switch signals are read from the switches on the air-conditioneroperation panel (step S1). Next, the sensor signals are read (step S2).Specifically, the inside air temperature (TR) detected by the inside airtemperature sensor 104, the outside air temperature (TAM) detected bythe outside air temperature sensor 105, the sunlight (TS) detected bythe sunlight sensor 106, the post-evaporator temperature (TE) detectedby the post-evaporator temperature sensor 107), the cooling watertemperature (TW) detected by the cooling water temperature sensor 108,and the discharge pressure (Pd) of the compressor 7 detected by therefrigerant pressure sensor 109.

Next, the target venting temperature (TAO) of the air vented to thepassenger compartment is calculated based on the following equation (1)stored in advance in the ROM (step S3):

TAO=Kset×Tset−KR×TR−KAM×TAM−KS×TS+C  (1)

Note that Tset is the temperature setting set by the temperature settingswitch 10, TR is the inside air temperature detected by the inside airtemperature sensor 104, TAM is the outside air temperature detected bythe outside air temperature sensor 105, and TS is the sunlight detectedby the sunlight sensor 106. Further, Kset, KR, KAM, and KS are gains,while C is a correction constant.

Next, it is judged if the air-conditioning mode is the cooler mode ornot. Specifically, it is judged if the target venting temperature (TAO)is below a predetermined temperature or if the cooler mode has been setto by the mode selection switch 100 (step S4). If the result of thejudgement is YES, the electromagnetic switch 8 is energized (ON), thefirst solenoid valve 23 opens, the second solenoid valve 24 closes, andthe refrigeration cycle apparatus 20 is operated by the refrigerationcycle circuit 21 (step S5).

Next, the electromagnetic coil 73 of the electromagnetic volume controlvalve 9 is energized (ON) (step S6), After this, the routine proceeds tothe processing of step S9. Accordingly, as shown in the illustration ofstep S6, the volume is controlled so that when the suction pressure (Ps)of the compressor 7 becomes lower, the discharge volume (Vc) of thecompressor 7 is made smaller, while when the suction pressure (Ps) ofthe compressor 7 becomes higher, the discharge volume (Vc) of thecompressor 7 is made larger.

Further, when the result of judgement of step S4 is NO, theelectromagnetic clutch 8 is energized (ON), the first solenoid valve 23closes, the second solenoid valve 24 opens, and the refrigeration cycleapparatus 20 is operated by the hot gas heater circuit 22 (step S7).Next, the electromagnetic coil 73 of the electromagnetic volume controlvalve 9 is deenergized (OFF) (step S8). Accordingly, as shown in theillustration of step S8, the volume is controlled so that when thesuction pressure (Ps) of the compressor 7 becomes lower, the dischargevolume (Vc) of the compressor 7 is increased, while when the suctionpressure (Ps) of the compressor 7 becomes higher, the discharge volume(Vc) of the compressor 7 is made smaller.

Next, it is judged if the load is a cooling load or a heating load basedon the target venting temperature (TAO). The target post-evaporatortemperature (TEO) is decided from the cooling load or heating load.Specifically, it is calculated so that the higher the target ventingtemperature (TAO), the higher the target post-evaporator temperature(TEO) becomes (step S9). Next, the volume of the compressor 7 iscontrolled so that the actual post-evaporator temperature (TE) detectedby the post-evaporator temperature sensor 107 becomes equal to thetarget post-evaporator temperature (TEO) (step S10). Specifically, thecontrol current to the electromagnetic coil 69 of the electromagneticvolume control valve 9 is controlled. Next, the routine of FIG. 5 isgone through.

Next, the operation of the vehicular air-conditioning system of thepresent embodiment will be briefly explained based on FIG. 1 to FIG. 7.Here, FIG. 6 shows the state of operation of the electromagnetic volumecontrol valve at the time of the cooler mode, while FIG. 7 shows thestate of operation of the electromagnetic volume control valve at thetime of the heater mode.

When the actual post-evaporator temperature (TE) has become considerablyhigher than the target post-evaporator temperature (TEO), the controlcurrent flowing through the electromagnetic coil 69 of theelectromagnetic volume control valve 9 is made smaller and the settingof the suction pressure (Ps) of the compressor 7 is made smaller. Inthis case, the bellows 65 contracts and thereby the valve element 63slightly displaces and the opening degree of the communication port 61becomes smaller. Due to this, the discharge pressure (Pd) of thecompressor 7 has difficulty entering the pressure passage 59 and thecrank chamber pressure (Pc) becomes smaller. By the crank chamberpressure (Pc) becoming smaller, the inclination of the swash plate 41 ofthe compressor becomes larger, so the stroke of the piston 42 becomeslonger. As a result, the discharge pressure (Pd) of the compressor 7becomes higher, so the discharge volume (Vc) of the compressor 7 becomeslarger.

Further, when the actual post-evaporator temperature (TE) becomessubstantially equal to the target post-evaporator temperature (TEO), thecontrol current flowing through the electromagnetic coil 69 of theelectromagnetic volume control valve 9 is made larger and the setting ofthe suction pressure (Ps) of the compressor 7 is made larger. In thiscase, by the expansion of the bellows 65, the valve element 63 displacesby a large amount and the opening degree of the communication port 61becomes larger. Due to this, the discharge pressure (Pd) of thecompressor 9 enters the pressure passage 59 and the crank chamberpressure (Pc) becomes larger. Further, by the crank chamber pressure(Pc) becoming larger, the inclination of the swash plate 41 of thecompressor becomes smaller and therefore the stroke of the piston 42becomes shorter. As a result, since the discharge pressure (Pd) of thecompressor 7 becomes lower, the discharge pressure (Vc) of thecompressor 7 becomes smaller.

Further, when the air-conditioning mode is the cooler mode, theelectromagnetic clutch 8 is turned on, the first solenoid valve 23opens, and the second solenoid valve 24 closes. Accordingly, the hightemperature, high pressure gas phase refrigerant discharged from thecompressor 7 circulates through the refrigeration cycle circuit 21 andflows into the evaporator 6. The air sucked into the air-conditioningduct 2 is heat exchanged with the low temperature, low pressurerefrigerant and cooled by the evaporator 6 and then vented into thepassenger compartment. The passenger compartment is cooled by this.

When the air-conditioning mode is the cooler mode, the electromagneticcoil 73 of the electromagnetic volume control valve 9 is energized (ON),so as shown in FIG. 6, the valve elements 71 and 72 displace downward inthe figure against the force of the return spring 74, whereby thepressure passage 56 and the communication passage 62 are communicatedand the pressure passage 58 and the communication passage 60 arecommunicated. Therefore, the discharge pressure (Pd) is led to the valveelement 63, so the lower the suction pressure (Ps) of the compressor 7becomes, the more the valve element 63 displaces to the valve openingside, the larger the opening degree of the communication port 61becomes, and the higher the crank chamber pressure (Pc) of thecompressor 7 becomes.

Due to this, when the suction pressure (Ps) is a low pressure of lessthan a preset first predetermined pressure (for example, a gaugepressure of 2 kg/cm²), the valve element 63 opens, the crank chamberpressure (Pc) rises due to the discharge pressure (Pd), and thedischarge volume (Vc) of the compressor 7 is controlled to 5% volume.Further, when the suction pressure (Ps) is a high pressure of more thana preset second predetermined pressure (for example, a gauge pressure of2.1 kg/cm²), the valve element 63 closes fully, the crank chamberpressure (Pc) becomes equal to the suction. pressure (Ps), and thedischarge volume (Vc) of the compressor 7 is controlled to 100% volume.

Further, when the suction pressure (Ps) becomes higher than the firstpredetermined pressure and lower than the second predetermined pressure,the valve element 63 displaces to the valve-closing side. the crankchamber pressure (Pc) becomes higher than the suction pressure (Ps), thedischarge pressure (P) is approached, and therefore the discharge volume(Vc) of the compressor 7 changes (see step S6 in FIG. 5).

Further, when the air-conditioning mode is the heater mode, theelectromagnetic clutch 8 is turned on, the first solenoid valve 23closes, and the second solenoid valve 24 opens. Further, the hot watervalve 15 also opens. Accordingly, the high temperature, high pressuregas phase refrigerant discharged from the compressor 7 circulatesthrough the hot gas heater circuit 22 and flows into the evaporator 6.Further, the cooling water absorbing the exhaust heat of the engine Ecirculates through the cooling water circulation circuit 14 and flowsinto the heater core 5. Further, the air sucked into theair-conditioning duct 2 is heat exchanged by the high temperature, lowpressure refrigerant at the evaporator 6 and further is heat exchangedwith the high temperature cooling water at the heater core 5 to befurther heated then is vented into the passenger compartment. Thepassenger compartment is heated by this.

Since the electromagnetic coil 73 of the electromagnetic volume controlvalve 9 is deenergized (OFF), as shown in FIG. 6, the valve elements 71and 72 displace upward in the illustration due to the force of thereturn spring 74, so the pressure passage 55 and the communicationpassage 60 are communicated and the pressure passage 57 and thecommunication passage 62 are communicated. Therefore, the suctionpressure (Ps) is led to the valve element 63, so the lower the suctionpressure (Ps) of the compressor 7, the more the valve element 63displaces to the valve opening side, the opening degree of thecommunication port 61 becomes larger, and the crank chamber pressure(Pc) of the compressor 7 becomes lower.

Due to this, when the suction pressure (Ps) is a low pressure of lessthan the preset first predetermined pressure (for example, a gaugepressure 3 kg/cm²), the valve element 63 fully opens, the crank chamberpressure (Pc) becomes equal to the suction pressure (Ps), and thedischarge volume (Vc) of the compressor 7 is controlled to 100% volume.Further, when the suction pressure (Ps) is a high pressure of less thanthe preset second predetermined pressure (for example, a gauge pressureof 3.1 kg/cm²), the valve element 63 closes, the crank chamber pressure(Pc) becomes higher than the discharge pressure (Pd), and the dischargevolume (Vc) of the compressor 7 is controlled to 5% volume.

Further, when the suction pressure (Ps) is higher than the firstpredetermined pressure and lower than the second predetermined pressure,the valve element 63 displaces to the valve closing side, the crankchamber pressure (Pc) becomes higher than the suction pressure (Ps), andthe discharge pressure (Pd) is approached, whereby the discharge volume(Vc) of the compressor 7 changes (see step S8 in FIG. 5).

As explained above, the vehicular air-conditioning system controls theair cooling performance (cooling performance) of the evaporator 6, theair heating performance (auxiliary heating performance) of theevaporator 6, and the discharge pressure (Pd) of the compressor 7 to theoptimum values by adjusting the discharge volume of the compressor 7 bythe electromagnetic volume control valve 9 in accordance with thecooling load and the heating load without turning the electromagneticclutch 8 on and off. By this, there is no frequent repeated turning onand off of the compressor 7, so the compressor 7 does not fluctuatelargely in torque. Accordingly, the rotational speed of the engine E forbelt driving the compressor 7 does not fluctuate by a large degree, sothe acceleration performance or slope climbing performance and otherpower performance and drivability of the vehicle do not deteriorate.

When the air-conditioning mode is a cooler mode, due to the use of theelectromagnetic volume control valve 9 of the present embodiment, whenthe cooling load becomes smaller and the suction pressure (Ps) of thecompressor 7 becomes lower, the discharge volume (Vc) of the compressor7 becomes smaller. Accordingly, the cooling performance of theevaporator 6 falls, so it is possible to suppress the occurrence of anexcessive cooling capacity or the occurrence of frost in the evaporator6.

Further, when the air-conditioning mode is the heater mode, if theheating load is large, for example, at the time of startup of the hotwater type heating system 4 in a low temperature environment where theoutside air temperature is less than a predetermined temperature (forexample, 0° C.)., when low temperature air is sucked into the evaporator6, the low temperature air is heat exchanged at the evaporator 6, so thetemperature and pressure of the refrigerant fall. Due to this, thesuction pressure (Ps) of the compressor 7 falls. By using theelectromagnetic volume control valve 9 of the present embodiment,however, even if the suction pressure (Ps) of the compressor 7 falls,the discharge volume (Vc) of the compressor 7 becomes larger. Therefore,by the increase of the flow rate of the circulation of the refrigerantthrough the hot gas heater circuit 22, the flow rate of the refrigerantinto the evaporator 6 increases. Accordingly, even when the cooling heatload is large, a sufficient auxiliary heating performance can beexhibited.

When the air-conditioning mode is the heater mode, due to use of theelectromagnetic volume control valve 9 of the present embodiment, if theheating load becomes smaller and the suction pressure (Ps) of thecompressor 7 becomes higher, the discharge volume (Vc) of the compressor7 becomes smaller. Accordingly, the auxiliary heating performance at theevaporator 6 falls and the discharge pressure (Pd) of the compressor 7becomes lower. Due to this, it is possible to prevent the inside airtemperature from becoming higher than the temperature setting and theauxiliary heating capacity becoming excessive or the refrigerant pipingand other cycle parts (refrigeration equipment) used in therefrigeration cycle apparatus 20 from breaking down or breaking.

FIG. 8 shows a second embodiment of the present invention and shows thegeneral structure of the electromagnetic volume control valve housed inthe compressor.

The electromagnetic volume control valve 9 of the present embodiment isa simplification of the first embodiment. Fixed calibrated orifices 53 aand 53 b are arranged in the communication passages 62 a and 62 b in therefrigerant pressure circuit and the fixed calibrated orifice 53 c isarranged in the pressure passage 55. The electromagnetic volume controlvalve 9 energizes (turns ON) the electromagnetic coil 73 when theair-conditioning mode is the cooler mode. Accordingly, as shown in FIG.8, the valve element 71 displaces downward in the figure against theforce of the return spring 74 whereby the pressure passage 58 andcommunication passage 60 are communicated.

Due to this, when the suction pressure (Ps) is a high pressure of morethan a preset second predetermined pressure, the valve element 63displaces to the most closed side, the crank chamber pressure (Pc)becomes equal to the suction pressure (Ps), and the discharge volume(Vc) of the compressor 7 is controlled to 100% volume. Further, when thesuction pressure (Ps) is higher than the preset first predeterminedpressure and lower than the second predetermined pressure, the valveelement 63 displaces to the valve opening side, the crank chamberpressure (Pc) becomes higher than the suction pressure (Ps), and thedischarge pressure (Pd) is approached. Due to this, change is possibleso that the lower the suction pressure (Ps) becomes, the smaller thedischarge volume (Vc) of the compressor 7 becomes.

The electromagnetic volume control valve 8 deenergizes theelectromagnetic coil 73 (turns it OFF) when the air-conditioning mode isthe heater mode. Accordingly, the valve element 71 displaces upward inthe figure due to the force of the return spring 74, whereby thepressure passage 57 and the communication passage 62 a are communicated.Due to this, when the suction pressure (Ps) is a low pressure of lessthan the preset first predetermined pressure, the valve element 63 fullyopens, the crank chamber pressure (Pc) becomes equal to the suctionpressure (Ps), and the discharge volume (Vc) of the compressor 7 iscontrolled to 100% volume. Further, when the suction pressure (Ps) ishigher than the first predetermined pressure and is lower than thesecond predetermined pressure, the valve element 63 displaces to thevalve closing side, the crank chamber pressure (Pc) becomes higher thanthe suction pressure (Ps), and the discharge pressure (Pd) isapproached. Due to this, control is possible so that the higher thesuction pressure (Ps) becomes, the smaller the discharge volume (Vc) ofthe compressor 7 becomes.

FIG. 9 shows a third embodiment of the present invention and shows thegeneral structure of the electromagnetic volume control valve housed inthe compressor.

The electromagnetic volume control valve 9 of the present embodimentdirectly communicates the communication port 61 and the pressure passage58 by the communication passage 60, communicates the pressure passage 59and pressure passage 55 by the communication passages 62 a and 62 b, anddisposes the solenoid valve 75 in the communication passage 62 a andtherefore is configured to control the discharge volume (Vc) to 100%fixed volume when the air-conditioning mode is the heater mode.

Further, in the cooler mode, by deenergizing (turning OFF) the solenoidvalve 75 and causing the valve to close, the result is a refrigerantpressure circuit similar to the cooler mode of the first embodiment.Further, in the heater mode, by energizing (turning ON) the solenoidvalve 75 and causing the valve to open, it is configured to similarlycontrol the crank chamber pressure (Pc) to the suction pressure (Ps) atall times and thereby fix the discharge volume (Vc) to 100% volumeregardless of the level of the suction pressure (Ps).

FIG. 10 and FIG. 11 shows a fourth embodiment of the present invention.Figure shows a general structure of the electromagnetic type volumecontrol valve housed in the compressor.

The electromagnetic type volume control valve 9 of the presentembodiment provides a high pressure control valve for changing thedischarge volume (Vc) of the compressor 7 in parallel with the object ofprotecting the refrigerant piping and other cycle parts at the time ofhigh pressure of the discharge pressure (Pd) of the compressor 7 andsuppressing the fluctuations in the rotational speed of the engine Ewhen turning the compressor from ON to OFF in state.

Further, the refrigerant pressure circuit of the high pressure controlvalve 80 is provided with the pressure passage 81 for guiding thesuction pressure (Ps) of the compressor 7, the pressure passages 82 and83 for guiding the discharge pressure (Pd) of the compressor 7, apressure passage 84 for giving a crank chamber pressure (Pc) to thecrank chamber 52 of the compressor 7, and a communication passage 85 forcommunicating the pressure passages 83 and 84. Note that the pressurepassage 81 is. provided with a fixed calibrated orifice 81 a. Further,the opening degree of the communication port 85 is determined by thestopping position of the valve element 85. The stopping position of thevalve element 86 is determined by the displacement position of the rod87 and the bellows 88. Note that reference numeral 89 is return springfor returning the bellows to the initial position.

Next, the control of the compressor capacity by the air-conditioner ECU10 of the present embodiment will be briefly explained based on FIG. 10and FIG. 11. Here, FIG. 11 is a flow chart of the method of control ofthe discharge volume by the air-conditioner ECU 10.

The processing of step S10 of the flow chart of FIG. 5 in the firstembodiment is performed, then the high pressure control valve 80 is usedto control the discharge volume (Vc) (step S11). Due to this, when thedischarge pressure (Pd) given to the pressure passage 82, rises from thepreset working pressure of the bellows 88, the valve element 86 opensand opens the communication port 85 and the crank chamber pressure (Pc)rises.

Accordingly, when the discharge pressure (Pd) of the compressor 7 is alow pressure of less than a first predetermined pressure (for example, agauge pressure of 20 kg/cm²), the discharge volume (Vc) of thecompressor 7 is controlled to become 100% volume. Further, when thedischarge pressure (Pd) of the compressor 7 is higher than the firstpredetermined pressure and lower than a second predetermined pressure(for example, a gauge pressure of 22 kg/cm²), control is performed sothat the higher the discharge pressure (Pd), the smaller the dischargevolume (Vc) becomes. Further, when the discharge pressure (Pd) of thecompressor 7 is a high pressure of more than the second predeterminedpressure, the discharge volume (Vc) of the compressor 7 is controlled tofor example 5% volume.

FIG. 12 and FIG. 13 shows a fifth embodiment of the present invention.FIG. 12 shows the general structure of the electromagnetic type volumecontrol valve housed in the compressor, the switching control valve, andthe hot gas volume control valve.

The electromagnetic type volume control valve 9 of the presentembodiment is a variable volume control means for the cooler mode. Theelectromagnetic type volume control valve 9 is provided with a returnspring 91 for returning the plunger 64 to the initial position, thespring seat 92 of the return spring 91, and the adjustment cock 93 foradjusting the amount of displacement of the plunger 64. Further, insidethe bellows 65 is provided a return spring 94 for returning the bellows65 to the initial position.

Further, at the end of the valve body 95 of the electromagnetic typevolume control valve 9 is provided a cock 96 for setting the initialload of the return spring 94. Note that the valve body 95 is formed witha pressure passage 95 c for giving the crank chamber pressure (Pc) tothe crank chamber 52 of the compressor 7, a pressure passage 95 d forguiding the discharge pressure (Pd) of the compressor 7, and a pressurepassage 95 s for guiding the suction pressure (Ps) of the compressor 7.

The refrigerant pressure circuit communicating with the electromagnetictype volume control valve 9 has disposed in it a switching control valve98 for changing the stopping position of the valve element 97 betweenthe cooler mode and the heater mode and a variable volume control meansfor the heater mode, that is, the hot gas volume control valve 99. Theswitching control valve 98 has a valve element 97, an electromagneticcoil 97 a, and a return spring 97 b. The switching control valve 98 isformed with a communication passage 98 a for communicating with thepressure passage 95 d, a communication passage 98 b for communicatingwith the hot gas volume control valve 99, and a pressure passage 98 dfor guiding the discharge pressure (Pd) of the compressor 7.

Further, the hot gas volume control valve 99 has a valve element 99 aand a bellows 99 b. The hot gas volume control valve 99 is formed with apressure passage 99 c for giving a crank chamber pressure (Pc) to thecrank chamber 52 of the compressor 7. Note that the pressure passage 99c communicates with the discharge port 51 through the crank chamber 52.Further, reference numeral 99 e is a return spring for returning thevalve element 99 a and bellows 99 b to the initial position.

In the present embodiment, when the air-conditioning mode is the coolermode, the electromagnetic coil 97 a of the switching control valve 98 isdeenergized (OFF) and the valve element 97 displaces upward in thefigure due to the force of the return spring 97 b to close thecommunication passage 98 b. Due to this, the discharge pressure (Pd) ofthe compressor 7 is guided to the pressure passage 95 d of theelectromagnetic type volume control valve 9.

Further, when the suction pressure (Ps) given to the pressure passage 95s is a high pressure of more than the second predetermined pressure (forexample, a gauge pressure of 2.1 kg/cm²), as shown in FIG. 13A, thebellows 65 contracts and the valve element 63 closes, whereby the crankchamber pressure (Pc) becomes equal to the suction pressure (Ps) and thedischarge volume (Vc) of the compressor 7 becomes 100% volume. Further,when the suction pressure (Ps) is a low pressure of less than the firstpredetermined pressure (for example, a gauge pressure of 2 kg/cm²), asshown in FIG. 13A, the bellows 65 expands and the valve element 63opens, whereby the crank chamber pressure (Pc) becomes equal to thedischarge pressure (Pd) and the discharge volume (Vc) of the compressor7 becomes 5% volume. Note that when the suction pressure (Ps) is higherthan the first predetermined pressure and lower than the secondpredetermined pressure, as shown in FIG. 13A, the discharge volume (Vc)of the compressor 7 is continuously changed from 5% volume to 100%volume the higher the suction pressure (Ps).

Further, when the air-conditioning mode is the heater mode, theelectromagnetic coil 97 a of the switching control valve 98 is energized(ON) and the valve element 97 displaces downward in the illustration toclose the communication passage 98 a. Due to this, the dischargepressure (Pd) of the compressor 7 is guided into the control chamber 99d of the hot gas volume control valve 99. Further, when the dischargepressure (Pd) given to the control chamber 99 d is a low pressure ofless than the first predetermined pressure (for example, a gaugepressure of 20 kg/cm²), as shown in FIG. 13B, the bellows 99 b expandsand the valve element 99 a closes, whereby the crank chamber pressure(Pc) becomes equal to the suction pressure (Ps) and the discharge volume(Vc) of the compressor 7 becomes 100% volume.

Further, when the discharge pressure (Pd) given to the control chamber99 d is a high pressure of more than the second predetermined pressure(for example, a gauge pressure of 22 kg/cm²), as shown in FIG. 13B, thebellows contracts and the valve element 99 a opens, whereby the crankchamber pressure (Pc) becomes equal to the discharge pressure (Pd) andthe discharge volume (Vc) of the compressor 7 becomes 5% volume. Notethat when the suction pressure (Ps) is higher than the firstpredetermined pressure and lower than the second predetermined pressure,as shown in FIG. 13B, the discharge volume (Vc) of the compressor 7 iscontinuously changed from 100% volume to 5% volume the higher thedischarge pressure (Pd).

FIG. 14 to FIG. 17 show a sixth embodiment of the present invention.FIG. 14 is a view of the overall structure of the vehicularair-conditioning system, FIG. 15 is a view of the general structure ofthe electromagnetic type volume control valve housed in the compressor,the switching control valve, and the high pressure control valve, andFIG. 16 is a view of the control system of a vehicular air-conditioningsystem.

The electromagnetic type volume control valve 9 of the presentembodiment is a variable volume control means for changing the settingof the discharge pressure (Pd) of the compressor 7 by the controlcurrent from the air-conditioner ECU 10 in the cooler mode and theheater mode as shown in FIG. 17A and thereby changing the dischargevolume (Vc) of the compressor 7. Further, in the present embodiment, theelectromagnetic type high pressure control valve 120 is provided insteadof the hot gas volume control valve 99 of the fifth embodiment.

The high pressure control valve 120 is a variable discharge pressuremeans which has a valve element 122 for changing the opening degree ofthe communication port 121 formed in the valve body 119 and sets thedischarge pressure (Pd) of the compressor 7 lower the closer the heaterventing temperature (TH) detected by the later mentioned heater ventingtemperature sensor 110 becomes to the target heater venting temperature(THO: for example, 50° C.).

The stopping position of the valve element 122 is configured to bedetermined by the displacement position of the plunger 123 and thebellows 124. That is, the plunger 123 and the bellows 124 are linkedwith the valve element 122 through the intermediate member 125 and therod 126. Further, the set position of the plunger 123 is configured tobe changed by the magnitude of the control current to theelectromagnetic coil 127.

Further, the valve body 119 is provided inside it with a return spring128 for returning the plunger 123 to the initial position. Further,inside the bellows 124 is provided a return spring 129 for returning thebellows 124 to its initial position. Further, the end of the valve body119 is provided with a cock 130 for setting the initial load of thereturn spring 129.

Further, the valve body 119 is formed with a pressure passage 131 givinga crank chamber pressure (Pc) to the crank chamber 52 of the compressor7 and pressure passages 132 and 133 guiding the discharge pressure (Pd)of the compressor 7 through the communication passage 98 b. Note thatthe pressure passage 131 and the pressure passage 132 are communicatedthrough a communication port 121 in the valve body 119. Accordingly, thehigh pressure control valve 120 is structured to send the dischargepressure (Pd) of the compressor 7 to a crank chamber (control pressurechamber) 52. The change of the opening degree of the valve element 122is determined by the expansion and contraction of the bellows 124 andthe balance of the force of the plunger 123 in accordance with thecontrol current to the electromagnetic coil 127.

On the other hand, the air-conditioner ECU 10 for controlling thevarious air-conditioning means in the air-conditioning unit 1, forexample, the electromagnetic clutch 8, the electromagnetic type volumecontrol valve 9, the blower fan motor 12, the drive motor 17, theswitching control valve 98, the high-side pressure control valve 120,etc. receives as input switch signals from various switches such as amode selection switch 100, a temperature control lever 111, anair-conditioner switch 102, and a blower fan switch 103. Of these, thetemperature control lever 111 instructs the maximum cooling operation(MAX COOL) when operated to one extreme side and instructs the maximumheating operation (MAX HOT) when operated to the other extreme side.

Further, the air-conditioner ECU 10 receives as input sensor signalsfrom various types of sensors such as an inside air temperature sensor104, outside air temperature sensor 105, sunlight sensor 106,post-evaporator temperature sensor 107, cooling water temperature sensor108, refrigerant pressure sensor 109, and heater venting temperaturesensor 110. Of these, the heater venting temperature 110 corresponds tothe venting temperature sensing means of the present invention and is aheater venting temperature detecting means for detecting the temperatureof the air directly after passing through the heater core 5 (hereinafterreferred to as a heater venting temperature).

Next, the control of the compressor capacity by the air-conditioner ECU10 of the present embodiment will be simply explained based on FIG. 14to FIG. 17.

The processing of step S10 of the flow chart of FIG. 5 of the firstembodiment is performed, then for example the discharge volume (Vc) ofthe compressor 7 is controlled by for example feedback control (PIcontrol). Specifically, the control current (I) of the compressor 7serving as the target value for the control current to be supplied forthe energization and deenergization of the electromagnetic coil 97 ofthe switching control valve 98 and for the electromagnetic coil 69 ofthe electromagnetic type volume control valve 9 and the electromagneticcoil 127 of the high pressure control valve 120 is calculated(determined) (control current processing means).

Specifically, the control current (In) is calculated based on thefollowing equation (2) and equation (3):

En=TH−THO  (2)

In=In−1−Kp{(En−En−1)+(θ/Ti)×En}  (3)

Note that TH is the actual heater venting temperature detected by theheater venting temperature sensor 110, THO is a preset target heaterventing temperature (for example, 50° C.), Kp is a proportionalconstant, θ is a sampling period (for example, 1 second), Ti is anintegration period, En is a current temperature error, En−1 is aprevious temperature error, In is a current control current, and In−1 isa previous control current.

Here, when the driver turns on the ignition switch to start the engineand start the hot water type heating system 4, the cooling water passingthrough the cooling water circulation circuit 14 to cool the engine Eflows into the heater core 5 in the air-conditioning duct 2. Further,when the temperature control lever 111 is set to the MAX HOT position,the outside air temperature (TAM) is a temperature lower than apredetermined temperature (for example, −5° C.), and the heater ventingtemperature (TH) is a temperature lower than the target heater ventingtemperature (THO), until a predetermined time (for example, 5 minutes to15 minutes) passes from when the engine E starts (startup time), thecooling water temperature is low and the heating capacity of the heatercore 5 is insufficient.

Therefore, the first solenoid valve 23 is closed, the second solenoidvalve 24 is opened, the refrigeration cycle apparatus 20 is switchedfrom the refrigeration cycle circuit 21 to the hot gas heater circuit22, the electromagnetic clutch 8 is turned ON, and the compressor 7 isstarted so as to augment the heating capacity of the hot water typeheating system 4. At this time, since the air-conditioning mode is theheater mode, the electromagnetic coil 97 a of the switching controlvalve 98 is energized (ON) and the valve element 97 displaces downwardin the figure so the communication passage 98 a is closed. Due to this,the discharge pressure (Pd) of the compressor is guided to the pressurepassages 132 and 133 through the communication passage 98 b.

Further, when the discharge pressure (Pd) given to the pressure passage133 is a low pressure of less than a first predetermined pressure (forexample, a gauge pressure of 20 kg/cm²), as shown in FIG. 17B, thebellows 124 expands and the valve element 122 closes, whereby the crankchamber pressure (Pc) becomes equal to the suction pressure (Ps) and thedischarge volume (Vc) of the compressor 7 becomes a large 100% volume.

Further, when the discharge pressure (Pd) given to the pressure passage133 is a high pressure of more than a second predetermined pressure (forexample, a gauge pressure of 22 kg/cm²), as shown in FIG. 17B, thebellows 124 contracts and the valve element 122 opens, whereby the crankchamber pressure (Pc) becomes equal to the discharge pressure (Pd) andthe discharge volume (Vc) of the compressor 7 becomes a small 5% volume.

Further, when the suction pressure (Ps) given to the pressure passage133 is higher than the first predetermined pressure and lower than thesecond predetermined pressure, as shown in FIG. 17B, the dischargevolume (Vc) of the compressor 7 is continuously changed from 100% volumeto 5% volume the higher the discharge pressure (Pd).

Here, by changing the control current to the electromagnetic coil 127 ofthe high pressure control valve 120 based on equations (2) and (3), thedischarge pressure (Pd) of the compressor 7 is set lower as shown by thearrow mark of FIG. 17B the close the heater venting temperature (TH)detected by the heater venting temperature sensor 110 becomes to thetarget heater venting temperature (THO; for example 50° C.).

Due to this, even right after the engine E is started when thetemperature of the cooling water, which had been low, rises, the heatingload becomes smaller, and the discharge pressure (Pd) of the compressor7 becomes lower, it is possible to make the discharge volume (Vc) of thecompressor 7 further smaller. Due to this, when the heater ventingtemperature (TH) approaches the target heater venting temperature (THO),the rotational speed of the engine E belt driving the compressor 7through the electromagnetic clutch 8 becomes the minimum necessary.Further, the flow rate of the refrigerant flowing inside the evaporator6 also falls. At this time, the air flowing inside the air-conditioningduct 2 is partially heated when passing through the evaporator 6 and isfully heated when passing through the heater core 5, whereby the ventingtemperature of the air becomes the optimum venting temperature and it ispossible to prevent the heating capacity of the evaporator 6 frombecoming excessive.

In the sixth embodiment, as the venting temperature detecting means, usewas made of the heater venting temperature sensor 110, but it is alsopossible to use the cooling water temperature sensor 108 as the ventingtemperature detecting means. That is, it is also possible to set thedischarge pressure (Pd) of the compressor 7 lower the closer the coolingwater temperature (TW) to the target cooling water temperature (TWO: forexample, 80° C.)

FIG. 18 shows a seventh embodiment of the application of the presentinvention to a refrigeration cycle apparatus in a vehicularair-conditioning system. The compressor 7 is driven through anelectromagnetic clutch 8 by the vehicle engine (not shown). At thedischarge side of the compressor 7 is connected a condenser 25 through afirst solenoid valve 23. At the outlet side of the condenser 25 isconnected a first pressure reducing means 27 through a check valve 18.The first pressure reducing means 27 is comprised of a capillary tube(fixed calibrated orifice) in this embodiment.

The outlet side of the first pressure reducing means 27 is connected toan evaporator 6. The outlet side of the evaporator 6 is connectedthrough an accumulator 28 to the suction side of the compressor 7. Onthe other hand, provision is made of a hot gas heater circuit 22 whichdirectly connects the discharge side of the compressor 7 to the inletside of the evaporator 6. The heater circuit 22 is provided with asecond solenoid valve 24 and second pressure reducing means 29 inseries. The second pressure reducing means 29 is comprised, in thisembodiment, by a constant pressure valve which opens when the dischargepressure of the compressor 7 reaches over a predetermined value.

The evaporator 6 is provided in the air-conditioning duct 2 of thevehicular air-conditioning system 2. The air blown from the centrifugalblower fan 3 (passenger compartment inside air or outside air) is cooledin the summer cooler mode. In the winter heater mode, the evaporator 6receives high temperature refrigerant gas (hot gas) flowing in from thehot gas heater circuit 22 and heats the air, so performs the function ofa radiator. In the air-conditioning duct 2 at the downstream air side ofthe evaporator 6 is arranged a hot water type heater core 5 which heatsthe vented air using the hot water from the vehicle engine as a heatsource. Air-conditioning air is vented from the vent (not shown)provided at the downstream side of the heater core 5 to the inside ofthe passenger compartment.

Next, explaining the portion of the accumulator 28, the key portion inthe present invention, in more detail, at the top surface of the tankportion 28 a of the accumulator 28 are provided an inlet passage 28 forintroducing the refrigerant from the outlet of the evaporator 6 and agas outlet passage 28 for guiding the gas phase refrigerant accumulatedat the upper region of the inside of the tank. At the bottom of the tankportion 28 a is provided a liquid outlet passage 28 for guiding theliquid phase refrigerant accumulated at the lower region of the insideof the tank.

The gas outlet passage 28 c and the liquid outlet passage 28 d merge andare connected to the suction side of the compressor 7. Further, theliquid outlet passage 28 d is provided with a valve mechanism 19 able tochange the opening degree (size of opening) of the calibrated orificepassage.

The valve mechanism 19 may specifically be comprised, as shown in FIG.19, of a first control valve 19 b which opens and closes a firstcalibrated orifice passage 19 a with a small opening degree (diameter)and a second control valve 19 d which opens and closes a secondcalibrated orifice passage 19 c with a large opening degree (diameter).Here the diameter of the first calibrated orifice passage 19 a is forexample ø1.2 and the diameter of the second calibrated orifice passage19 c is for example ø2.5.

The first and second control valves 19 b and 19 d may for example bycomprised of solenoid valves. These first and second control valves 19 band 19 d are opened and closed by energization under the control of theair-conditioner ECU 10 as shown in FIG. 20. In addition, theelectromagnetic clutch 8, first and second solenoid valves 23 and 24,blower fan 3, and the like are also actuated under the control of theair-conditioner ECU 10. The air-conditioner ECU 10 receives as inputsignals from a group of various sensors Se for control of theair-conditioning and operating switches Sw of the air-conditioneroperation panel as is well known.

Next, an explanation will be made of the operation of the seventhembodiment of the present invention. In the summer cooler mode, thefirst solenoid valve 23 is opened and the second solenoid valve 24 isclosed by the air-conditioner ECU 10. At the same time, the first andsecond control valves 19 b and 19 d are placed in the states shown inFIG. 19(a), that is, the first control valve 19 b is opened and thesecond control valve 19 d is closed.

Accordingly, the electromagnetic clutch 8 is engaged and the compressor7 is driven by the vehicle engine, whereby the discharge gas phaserefrigerant of the compressor 7 passes through the open first solenoidvalve 23 and flows into the condenser 25. In the condenser 25, therefrigerant is cooled and condensed by the outside air blown from thecooling blower fan 16. Further, the condensed liquid phase refrigerantpasses through the check valve 18 and is reduced in pressure by thefirst pressure reducing means 27 to become a low temperature, lowpressure two-phase liquid and gaseous state.

Next, the low pressure refrigerant flows into the evaporator 6, absorbsheat from the air-conditioning air blown by the blower fan 3, andevaporates. The air-conditioning air cooled by the evaporator 6 isvented into the vehicle passenger compartment to cool the passengercompartment. The refrigerant passing through the evaporator 6 flows fromthe inlet passage 28 b of the accumulator 28 to the inside of the tankportion 28 a. In the tank portion 28 a, the gas phase refrigerant andthe liquid phase refrigerant are separated by the difference in theirspecific gravities. The gas phase refrigerant accumulates in the upperregion of the tank portion 28 a. The gas phase refrigerant passesthrough the gas outlet passage 28 c and is sucked into the compressor 7.

Further, in the liquid outlet passage 28 d of the accumulator 28, sincethe first control valve 19 b is closed, the liquid phase refrigerantaccumulated at the lower part of the tank portion 28 a (includinglubrication oil) passes through the small diameter first calibratedorifice portion 19 a and is sucked into the compressor 7. Here, thediameter of the first calibrated orifice passage 19 a is set to theminimum value (for example, ø1.2) required for securing an amount ofreturn oil required for lubrication of the compressor 7, whereby theincrease in the amount of oil recirculated to the cycle can besuppressed and the cooling capacity improved. Due to the suppression ofthe amount of liquid phase refrigerant sucked into the compressor 7, theenergy consumption of the compressor can also be reduced.

In the winter heater mode, the first solenoid valve 23 is closed, thesecond solenoid valve 24 is opened, and the hot gas heater circuit 22 isopened up by the air-conditioner ECU 10. At the same time, the first andsecond control valves 19 b and 19 d enter the states shown in FIG. 19B,that is, the first control valve 19 b is closed and the second controlvalve 19 d is opened. These open and closed states of the valves may beswitched between by judgement by the air-conditioner ECU 10 of theconditions where the maximum heating state is required and thetemperature of the hot water flowing into the heater core 5 falls belowa predetermined value.

Further, as explained above, if the open and closed states of the valvesare changed, the high temperature discharge gas refrigerant (superheatedgas phase refrigerant) of the compressor 7 passes through the open statesecond solenoid valve 24 and is reduced in pressure by the secondpressure reducing means 29, then the reduced pressure superheated gasphase refrigerant releases its heat by the venting air of the evaporator6 and heats the vented air. Further, the gas phase refrigerant releasingheat in the evaporator 6 flows from the inlet passage 28 b of theaccumulator 28 into the tank portion 28 a, passes through the gas outletpassage 28 c, and is sucked into the compressor 7.

The amount of heat discharged from the gas phase refrigerant at theevaporator 6, however, corresponds to the amount of compression work ofthe compressor 7, so to increase the amount of heat released at theevaporator 6, it is necessary to increase the amount of compression workof the compressor 7.

Therefore, in the heater mode, in the accumulator 28, the second controlvalve 19 d is opened and the liquid phase refrigerant (includinglubrication oil) passing through the large diameter second calibratedorifice portion 28 a is sucked into the compressor 7. Due to this,compared with the cooler mode, in the heater mode, the amount of liquidphase refrigerant sucked from the accumulator 28 into the compressor 7increases. The amount of compression work can therefore be increased andthe heating capacity improved.

Note that as the second pressure reducing means 29, in the presentembodiment, use is made of a constant pressure valve which opens whenthe discharge pressure of the compressor 7 rises above a predeterminedvalue. Further, the check valve 18 prevents the gas phase refrigerantfrom the hot gas heater circuit 22 from flowing back into the condenser25 and the refrigerant accumulating in condenser 25 during the heatermode.

FIG. 21 is a graph showing the rotational speed Nc of the compressoralong the ordinate and the heating capacity along the abscissa and showsthe results of experiments conducted by the present inventors. As willbe understood from the graph of FIG. 21, the heating capacity can beincreased along with an increase of the diameter of the calibratedorifice passage of the liquid outlet passage 28 d of the accumulator 28.

FIG. 22 shows an eighth embodiment where two calibrated orifice passages19 a and 19 c with identical opening degrees (identical diameters) areprovided in parallel in the valve mechanism 19, a control valve 19 b isprovided at only one calibrated orifice passage 19 a, the control valve19 b is closed in the cooler mode, and the control valve 19 b is openedin the heater mode. Due to this, in the cooler mode, the liquid phaserefrigerant passes through just the one calibrated orifice passage 19 cto be sucked into the compressor 7, while in the heater mode, the liquidphase refrigerant passes through the parallel circuit of the twocalibrated orifice passages 19 a and 19 c and is sucked into thecompressor 7, so the amount of liquid phase refrigerant sucked into thecompressor 7 in the heater mode can be increased. The rest of theembodiment is the same as the seventh embodiment.

FIG. 23 shows a ninth embodiment in which as the valve mechanism 19, useis made of a solenoid valve mechanism which can continuously control theopening degree of the calibrated orifice passage 19 a of the liquidoutlet passage 28 d. The valve mechanism 19 is provided with aball-shaped valve element 19 e for adjusting the opening degree of thecalibrated orifice passage 19 a, a spring 19 f giving a spring force tothe valve element 19 e in the closing direction, and an electromagneticcoil 19 g giving an electromagnetic attraction force to the valveelement 19 e. By continuously changing the current to theelectromagnetic coil 19 g, it is possible to continuously adjust theopening degree of the calibrated orifice passage 19 a.

FIG. 24 is a graph illustrating the relationship between the current tothe electromagnetic coil 19 g due to the valve mechanism 19 and theopening degree of the calibrated orifice passage 19 a. By taking note ofthe operating characteristic shown in FIG. 24 and the operatingcharacteristic shown in FIG. 21 and controlling the current to theelectromagnetic coil 19 g in accordance with the operating conditions inthe heater mode by the air-conditioner ECU 10 (FIG. 20), it is possibleto change the opening degree of the calibrated orifice passage 19 a inaccordance with the heating capacity required at the time of the heatermode.

Explaining this more in more detail, by detecting the outside airtemperature by the group of sensors Se and increasing the current to theelectromagnetic coil 19 g as the outside air temperature falls, it ispossible to increase the heating capacity in accordance with a fall inthe outside air temperature.

Further, by detecting the passenger compartment temperature (inside airtemperature) instead of the outside air temperature by the group ofsensors Se and increasing the current to the electromagnetic coil 19 gin accordance with a fall in the inside air temperature, it is possibleto increase the heating capacity in accordance with a fall in the insideair temperature.

Further, by calculating the necessary temperature of the air vented(TAO) into the passenger compartment based on the outside airtemperature, inside air temperature, and temperature setting of thedriver or passengers by the air-conditioner ECU 10 and increasing thecurrent to the electromagnetic coil 19 g as the required temperature ofthe vented air (TAO) rises in the heater mode, it is possible toincrease the heating capacity in accordance with a rise in the requiredtemperature of the vented air (TAO).

Further, since there is a correlation between the discharge sidepressure of the compressor 7, that is, the high-side pressure, and thetemperature of the gas phase refrigerant flowing into the evaporator 6,by increasing the current to the electromagnetic coil 19 g as thehigh-side pressure falls, it is possible to prevent in advance the fallin the heating capacity caused by the fall in the high-side pressure.

FIG. 25 shows a 10th embodiment where the valve mechanism 19 of theninth embodiment is made integral with the accumulator 28. In FIG. 25,an inlet passage 28 b is formed at the top surface of the cylindricaltank portion 28 a of the accumulator 28. The inlet passage 28 b iscommunicated with the inside of the tank through a hole (not shown)formed in the top surface of the tank portion 28 a. Since an umbrellashaped guide member 28 e is affixed to the inside wall of the topsurface of the tank portion 28 a, the refrigerant from the inlet passage28 b flows into the tank along the outer surface of the guide member 28e.

On the other hand, an electromagnetic drive portion 19 h of the valvemechanism 19 is arranged at the outside of the bottom of the tankportion 28 a. This electromagnetic drive portion 19 h is provided with afixed magnetic pole member 19 i and a movable magnetic pole member(plunger) 19 j which moves along the fixed magnetic pole member 19 i dueto the electromagnetic attraction force of the electromagnetic coil 19g. This movable magnetic pole member 19 j is connected through a shaft19 k and piston-shaped linkage 19 m to the spherical valve element 19 e.Therefore, the movable magnetic pole member 19 j and the ball-shapedvalve element 19 e displace integrally in the vertical direction of FIG.25.

At the inside of the bottom of the tank portion 28 a is formed a liquidoutlet passage 28 d for carrying the liquid refrigerant. A calibratedorifice passage 19 a is formed in the middle of the liquid outletpassage 28 d. The opening degree of the calibrated orifice passage 19 ais made to be able to be continuously changed by the ball-shaped valveelement 19 e. The outlet side of the calibrated orifice passage 19 a iscommunicated with the inside of the bottom of the cylindrical member 28f through a communication hole 19 p of a holding case 19 n of a spring19 f giving a spring force to the valve element 19 d force in adirection closing the valve.

The cylindrical member 28 f is arranged so that the center portion ofthe inside of the tank portion 28 a extends in the vertical direction.At the center of the inside of the cylindrical member 28 f isconcentrically arranged a refrigerant outlet pipe 28 g. In this way, thegas phase refrigerant accumulated in the upper region of the tankportion 28 g passes from the opening at the lower end of the refrigerantoutlet pipe 28 g into the pipe 28 g and flows out to the outside of theaccumulator 28.

Further, the liquid phase refrigerant accumulated at the lower region ofthe tank portion 28 a passes through the liquid outlet passage 28 d, thecalibrated orifice passage 19 a, the communication hole 19 p, etc. toflow into the inside of the bottom of the cylindrical member 28 f, whereit is mixed with the gas phase refrigerant and is sucked into therefrigerant outlet pipe 28 g. Accordingly, in this embodiment, the gasoutlet passage 28 c is comprised of the inside space of the cylindricalmember 28 f.

In the 10th embodiment as well, however, since the amount ofdisplacement of the ball-shaped valve element 19 e can be continuouslyadjusted by the control of the current to the electromagnetic coil 19 g,in the same way as in the ninth embodiment, it is possible to controlthe opening degree of the calibrated orifice passage 19 a to the optimalvalue in accordance with the operating conditions in the heater mode.

FIG. 26 shows an 11th embodiment in which a receiver 26 is arrangedbetween the condenser 25 and check valve 18. In the receiver 26, theliquid and gas phases of the refrigerant condensed at the condenser 25are separated. The liquid phase refrigerant is accumulated, while thegas phase refrigerant is guided to the check valve 18 side. The presentinvention can be similarly applied to a refrigeration cycle apparatushaving such a receiver 26.

As the first pressure reducing means 27, in this embodiment, use is madeof a temperature type expansion valve which adjusts the opening degree(refrigerant flow rate) so that the superheating of the refrigerant atthe outlet of the evaporator 6 is maintained at a predetermined value.Reference numeral 27 a is a temperature sensing member for sensing thetemperature of the refrigerant at the outlet of the evaporator 6.

FIG. 27 shows a 12th embodiment. In the 12th embodiment, note is takenof the fact that in a refrigeration cycle apparatus having the receiver26 of FIG. 26, the surface of the liquid phase refrigerant in theaccumulator 28 changes by a large degree between the cooler mode and theheater mode and the opening degrees of the calibrated orifice passages19 a and 19 c of the liquid outlet passage 28 d of the accumulator 28are switched accordingly.

That is, in a refrigeration cycle apparatus having the receiver 26,since, as the first pressure reducing means 27, use is made of atemperature type expansion valve which adjusts the valve opening degree(refrigerant flow rate) so that the superheating of the refrigerant atthe outlet of the evaporator 6 is maintained at a predetermined value,during the cooler mode, the refrigerant at the outlet of the evaporator6 is maintained in the superheated gas state having superheating at alltimes by the temperature type expansion valve 27. Therefore, only thelubrication oil in the cycle is accumulated in the accumulator 28.

As opposed to this, in the heater mode, the superheating of therefrigerant at the outlet of the evaporator is not controlled by thetemperature type expansion valve 27, the high temperature refrigerantgas (hot gas) from the hot gas heater circuit 22 flows directly into theevaporator 6, and part of the high temperature refrigerant gas condensesat the evaporator 6, so both the liquid refrigerant and the lubricationoil accumulate in the accumulator 28.

As a result, the height of the liquid level in the accumulator 28becomes higher in the heater mode and becomes lower in the cooler mode.

Therefore, in the 12th embodiment, the small opening cooling use firstcalibrated orifice passage 19 a is arranged at the downward side in thevertical direction of the accumulator 28 (near bottom of accumulator 28)and the large opening heating use second calibrated orifice passage isarranged at the upward side in the vertical direction of the accumulator28.

FIG. 27 illustrates the specific structure of the accumulator 28according to the 12th embodiment. Parts the same as or equivalent tothose in the accumulator 28 of FIG. 25 are given the same referencenumerals and explanations thereof are omitted.

In the accumulator 28 of the 12th embodiment as well, provision is madeof a cylindrical member 28 f which extends in the vertical direction inthe center of the inside of the tank portion 28 a and a refrigerantoutlet pipe 28 g arranged concentrically with the center of the insideof the cylindrical member 28 f. By this, the gas phase refrigerantaccumulated at the top region in the tank portion 28 a passes throughthe inside space of the cylindrical member 28 f as shown by the arrow A,that is, through the gas outlet passage 28 c, then passes from the loweropening of the refrigerant outlet pipe 28 g through the pipe 28 g andflows to the outside of the accumulator 28.

On the other hand, at the lower end of the cylindrical member 28 f isconnected a cylindrical cap member 28 h with a bottom portion so as toclose the opening of the bottom portion of the cylindrical member 28 f.At a position near the bottom portion of the cap member 28 h (forexample, a position about 10 mm higher than the bottom surface of thetank portion 28 a) is provided the above small opening cooling use firstcalibrated orifice passage 19 a. The diameter of the first calibratedorifice passage 19 a is for example ø1.0.

On the other hand, the large opening heating use second calibratedorifice passage 19 c is provided passing through the engaging portion ofthe cylindrical member 28 f and the cap member 28 h exactly apredetermined dimension (for example, about 20 to 30 mm) above the firstcalibrated orifice passage 19 a. The diameter of the second calibratedorifice passage 19 c is for example ø2.3.

Note that in FIG. 27, reference numeral 28 i shows a support stay whichis arranged between the outer circumference of the cylindrical member 28f and the inner wall of the tank portion 28 a and stably supports thecylindrical member 28 f. A plurality of (four) support stays 28 i arearranged to extend in a radial manner from the outer circumference ofthe cylindrical member 28 f. Reference numeral 28 j is a desiccant forabsorbing the moisture in the cycle.

According to the 12th embodiment, in the cooler mode, just thelubrication oil in the cycle is accumulated in the accumulator 28 due tothe control of the superheating of the outlet refrigerant of theevaporator by the temperature type expansion valve 27 of FIG. 26, so theliquid level in the accumulator 28 falls below the second calibratedorifice passage 19 c. L₁ of FIG. 27 shows the liquid level in the coolermode.

Therefore, in the cooler mode, the lubrication oil passing through onlythe first calibrated orifice passage 19 a positioned below the liquidlevel L₁ and near the bottom of the accumulator 28 is sucked into thecylindrical member 28 f and returned to the suction side of thecompressor 7.

As opposed to this, in the heater mode, both the liquid phaserefrigerant and the lubrication oil accumulate in the accumulator 28 asmentioned above. The height of the liquid level in the accumulatorbecomes sufficiently high compared with the cooler mode and rises toabove the second calibrated orifice passage 19 c. L₂ of FIG. 27 showsthe liquid level in the heater mode.

Therefore, in the heater mode, the liquid phase refrigerant and thelubrication oil in the accumulator 28 pass through both of the firstcalibrated orifice passage 19 a and the second calibrated orificepassage 19 c and are sucked to the inside of the cylindrical member 28and returned to the suction side of the compressor 7. Therefore, it ispossible to increase the amount of liquid phase refrigerant andlubrication oil sucked into the compressor 7 and improve the heatingcapacity.

Further, according to the 12th embodiment, since use is made of the factthat the level of the liquid phase refrigerant in the accumulator 28changes by a large amount between the cooler mode and the heater modeand the opening degrees of the calibrated orifice passages 19 a and 19 cin the cooler mode and heater mode are switched, it is possible toeliminate the valve mechanism for changing the opening degrees andthereby simplify the configuration. Further, since the first and secondcalibrated orifice passages 19 a and 19 c are housed in the accumulator28 and there is no need to provide any additional mechanisms at theoutside of the accumulator, the accumulator 28 can be made smaller inshape.

Note that in the above embodiment, provision was made of a secondpressure reducing means 29 in the hot gas heater circuit 22 for theheater mode and the second pressure reducing means 29 was used to reducethe pressure of the gas phase refrigerant discharged from the compressorand make the refrigerant flow into the evaporator 6, the second pressurereducing means 29 of the hot gas heater circuit 22 may be abolished andthe outlet portion of the hot gas heater circuit 22 may be connected tothe upstream portion of the first pressure reducing means 27 and the gasphase refrigerant passing through the hot gas heater circuit 22 reducedin pressure by the first pressure reducing means 27 and then made toflow into the evaporator 6.

Further, while use was made of the first and second solenoid valves 23and 24 as the switching means for switching the gas discharged from thecompressor between the condenser 25 side passage and the hot gas heatercircuit 22 side, it is of course also possible to replace these by asingle three-way switching valve.

Further, in the 12th embodiment, the opening degree of the firstcalibrated orifice passage was made small and the opening degree of thesecond calibrated orifice passage 19 c was made large, but since theliquid phase refrigerant and lubrication oil flow in from both of thefirst and second calibrated orifice passages 19 a and 19 c, the openingdegrees of the first and second calibrated orifice passages may also bemade the same.

To augment the heating capacity of the heater core 5, further, whenswitching the refrigeration cycle apparatus 20 to the hot gas heatercircuit 22, the control current to the electromagnetic coil 69 is set toOA and the setting of the suction pressure (Ps) is made for example −1kg/cm²G for use of the electromagnetic volume control valve 9.

Further, even if the control pressure (Pc) is controlled to fall themost in this way, when the temperature of the outside air where the hotgas heater circuit 22 is to be used is in the extremely cold region ofless than −10° C., for example, as shown by the broken line in the timechart of FIG. 28, the saturation pressure of the refrigerant becomesless than 1 kg/cm²G and the pressure difference (between the high-sidepressure and the low-side pressure of the refrigeration cycle apparatus20, which is a factor increasing the discharge volume discharged fromthe discharge port of the compressor 7, cannot be obtained, so thedischarge volume will never become large.

Therefore, in the case of a vehicular use air-conditioning system whichis controlled to turn the blower fan motor 12 off until the coolingwater temperature (TW) of the engine E rises above a predeterminedtemperature (for example, 40° C.) at the startup time of the heater modeto prevent cold air from being vented into the passenger compartment,when mounting an engine with a small exhaust heat, the cooling watertemperature (TW) will not rise above the predetermined temperature (forexample, 40° C.) TWa in an extremely cold region where the temperatureof the outside air (TAM) is less than −30° C., so the centrifugal blowerfan 13 will end up not operating and the passenger compartment will notbe able to be heated ever.

Therefore, in a 13th embodiment of the present invention, at the time ofstartup of the auxiliary heating operation motor for augmenting theheating capacity of the heater core 5 of the hot water type heatingsystem, that is, at the time of startup of the compressor 7 at the timeof operation of the refrigeration cycle apparatus 20 in the hot gasheater circuit 22, the first and second solenoid valves 23 and 24 areboth closed and the high-side pressure of the refrigeration cycleapparatus 20, that is, the discharge pressure (Pd) of the discharge portof the compressor 7, is made easier to rise above 2 kg/cm²G so asthereby to increase the discharge volume of the compressor 7. Next, thesecond solenoid valve 24 is opened to constitute the hot gas heatercircuit 22. The changes in the discharge pressure (Pd) and the suctionpressure (Ps) of the compressor 7 are shown by the solid lines in thetime chart of FIG. 28.

Accordingly, with the second solenoid valve 24 left open, there is nodifference between the discharge pressure (Pd) and the suction pressure(Ps) and the discharge volume remains at a minimum. When the compressor7 is started up, however, the second solenoid valve 25 is closed untilpredetermined conditions are satisfied, whereby the discharge pressure(Pd) suddenly rises and therefore the discharge volume of the compressor7 becomes larger, it was learned.

Therefore, at the time of startup of the auxiliary heating operationmotor for augmenting the heating capacity of the heater core 5 of thehot water type heating system, that is, at the time of startup of thecompressor 7 at the time of operation of the refrigeration cycleapparatus 20 in the hot gas heater circuit 22, the conditions forclosing the second solenoid valve 24 (predetermined conditions) are forexample that the high-side pressure (discharge pressure) of therefrigeration cycle apparatus 20 detected by a refrigerant pressuresensor 109 be less than 2 kg/cm²G, the suction temperature of the airsucked into the evaporator 6 (evaporation suction temperature) be lessthan 0° C., etc. Note that an evaporator suction temperature of lessthan 0° C. means that when the suction port mode is the inside aircirculation mode, the inside air temperature (TR) detected by an insideair temperature sensor 104 is less than 0° C. and that when the suctionport mode is the outside air introduction mode, the outside airtemperature (TAM) detected by an outside air temperature sensor 105 isless than 0° C.

Further, the conditions for opening the second solenoid valve 24 afterthe startup of the auxiliary heating operation mode, that is, after thestartup of the compressor 7 in the case of operating the refrigerationcycle apparatus 20 by the hot gas heater circuit 22, are for examplethat the high-side pressure of the refrigeration cycle apparatus 20 riseto more than 2 kg/cm²G, about 10 seconds pass after startup of thecompressor 7, etc.

As explained above, at the time of startup of the auxiliary heatingoperation mode, even if the outside air temperature (TAM) is less than0° C. (in particular less than −20° C.), the air-conditioning unit 1 ofthe 13th embodiment closes the second solenoid valve 24 after thestartup of the compressor 7 until predetermined conditions aresatisfied, whereby it is possible to raise the discharge pressure (Pd)of the compressor 7 and therefore increase the difference between thepressure levels of thee refrigeration cycle apparatus 20.

Due to this, even if an external variable volume compressor isincorporated in the refrigeration cycle apparatus 20, it is possible toincrease the discharge volume of the compressor 7, so it is possible tosend a sufficient flow rate of refrigerant to the evaporator 6. Due tothis, even if the outside air temperature (TAM) is less than 0° C., theheating capacity of the evaporator 6 can be improved, so in therefrigeration cycle apparatus 20 of the 13th embodiment, it is possibleto sufficiently bring out the auxiliary heating performance foraugmenting the heating capacity of the heater core 5.

Further, the air-conditioning unit 1 of the 13th embodiment can raisethe heat radiation temperature of the evaporator 6 immediately afterstartup of the engine E at the time of startup of the auxiliary heatingoperation mode, so the surface temperature of the heater core 5 placedin the air-conditioning duct 2 near the evaporator 6 rises and thetemperature of the cooling water circulating through the heater core 5rises faster. Further, since the compressor 7 is belt driven by theengine E through an electromagnetic clutch 8 at the time of theauxiliary heater mode, the compressor 7 increases the drive load of theengine E. Due to this, the amount of exhaust heat of the engine becomeslarger, so the temperature of the cooling water circulating in thecooling water circulation circuit 14 rises faster.

Due to this, since the temperature of the cooling water rises above thepredetermined temperature (for example, 40° C.) Ta, even whencontrolling the apparatus to delay the start of the blower fan, thecentrifugal blower fan 13 immediately starts turning and the passengercompartment can be quickly heated.

FIG. 29 to FIG. 31 show a 14th embodiment of the present invention. FIG.29 shows a refrigeration cycle apparatus of a vehicular air-conditioningsystem, FIG. 30 shows a variable throttling valve provided in therefrigeration cycle apparatus, and FIG. 31 is a graph of the openingdegree of the variable throttling valve with respect to the high-sidepressure of the refrigeration cycle apparatus.

In the refrigeration cycle apparatus 20 of the 14th embodiment, thefixed calibrated orifice 29 of the 13th embodiment is changed to thevariable throttling valve 140. This variable throttling valve 140corresponds to the refrigerant passage throttling means of the presentinvention and is comprised of a valve housing 143 formed with an orifice142 in the middle of the communication passage 141 communicating withthe refrigerant passage leading the refrigerant from the second solenoidvalve 24 to the evaporator 6, a ball-shaped valve element 144 arrangedto be able to displace back and forth in the valve housing 143, adiaphragm 147 driving the valve element 144 through an operating rod 145and stopper 146, and an adjustment spring 149 which enables adjustmentof the opening pressure of the valve element 144 by an adjustment screw148.

Among these, the valve element 144 adjusts the Opening degree of theorifice 142 and is provided with a spring seat 150 with which theadjustment spring 149 abuts at the bottom of the illustration. Thediaphragm 147 corresponds to the valve element driving means of thepresent invention and is housed in the housing 151. Further, thehigh-side pressure of the refrigeration cycle apparatus 20 acts in thepressure chamber 152 formed by the diaphragm 147 and the housing 151.

The variable throttling valve 140 of the 14th embodiment, due to theabove configuration, receives the high-side pressure of therefrigeration cycle apparatus 20 in the pressure chamber 101, closeswhen the high-side pressure is less than 2 kg/cm²G as shown in the graphof FIG. 31, and opens widely along with a rise in the high-sidepressure, whereby it is possible to increase the discharge volume of thecompressor 7 even if the outside air temperature (TAM) is less than −20°C. at the time of startup of the auxiliary heater mode.

FIG. 32 shows a 15th embodiment of the present invention and illustratesa differential pressure valve built in the refrigeration cycleapparatus.

In the 15th embodiment, a differential pressure valve 160 is provided inthe middle of the refrigerant passage from the discharge port of thecompressor 7 to the inlet of the fixed throttle portion 29 in the hotgas heater circuit 22 of the 13th embodiment. This differential pressurevalve 160 is comprised of a valve body 161, a valve element 162 arrangedto be able to displace back and forth in the valve body 161, and anadjustment spring 164 which enables adjustment of the opening pressureof the valve element 162 by an adjustment screw 163. Further, the valveelement 162 is formed with a hook like stopper 167. An O-ring 168 is fitover the outer circumference of the valve element 162.

The differential pressure valve 160 of the 15th embodiment, beingarranged in the middle of the refrigerant passage from the dischargeport of the compressor 7 to the inlet of the fixed calibrated orifice29, fully closes the orifice 165 at the time of startup of the auxiliaryheater mode to increase the discharge volume of the compressor 7. Whenthe pressure difference before and after the valve element 162 becomeslarge, that is, when the high-side pressure rises higher than theopening pressure of the valve element 162 (for example, 2 kg/cm²G), thevalve element 162 opens and a hot gas heater circuit 20 is formed in therefrigeration cycle apparatus 20.

In the present embodiments, the outside variable volume compressor wasconstituted by the refrigerant compressor 7, electromagnetic clutch 8,electromagnetic volume control valve 9, etc., but the outside variablevolume control compressor may also be configured by the refrigerantcompressor 7, the electromagnetic type volume control valve 9, etc.without provision of the electromagnetic clutch 8 or other clutch means.In this case, the refrigerant compressor 7 is made to be directly drivenby the internal combustion engine.

In the present embodiments, further, at the time of startup of theauxiliary heater mode, the hot gas heater circuit 22 was fully closedafter startup of the refrigerant compressor 7 until predeterminedconditions were satisfied and the various valve devices were opened toform the hot gas heater circuit 22 after these predetermined conditionswere satisfied, but it is also Ha possible to throttle down (not fullyclose) the sectional area of the refrigerant passage constituting thehot gas heater circuit 22 compared with that of normal operation afterstartup of the compressor 7 until predetermined conditions weresatisfied when starting up the auxiliary heater mode.

In the present embodiments, further, the example was shown of use of anoutside variable volume type compressor provided with an electromagnetictype volume control valve 9 which increases the discharge volume of therefrigerant discharged from the discharge port of the compressor 7 whenthe suction pressure of the refrigerant sucked into the suction port ofthe compressor 7 becomes high, but use may also be made of a variablevolume type compressor provided with a variable discharge volume meanswhich reduces the discharge volume of the refrigerant discharged fromthe discharge port of the refrigerant compressor when the suctionpressure of the refrigerant sucked into the suction port of therefrigerant compressor becomes high. Further, use may be made of avariable volume type compressor provided with a variable dischargevolume means which reduces the discharge volume of the refrigerantdischarged from the discharge port of the refrigerant compressor whenthe discharge pressure of the refrigerant discharged from the dischargeport of the refrigerant compressor becomes high.

In the above embodiments, the present invention was applied to therefrigeration cycle apparatus for a vehicular air-conditioning systemfor an automobile etc., but the present invention may also be applied tothe refrigeration cycle apparatus of an air-conditioning system of anaircraft, ship, railroad car, etc. Further, the present invention mayalso be applied to the refrigeration cycle apparatus of anair-conditioning system of a factory, store, house, etc.

While the invention has been described by reference to specificembodiments chosen for purposes of illustration, it should be apparentthat numerous modifications could be made thereto by those skilled inthe art without departing from the basic concept and scope of theinvention.

What is claimed is:
 1. A refrigeration cycle apparatus comprising: (a) arefrigerant compressor driven in rotation by an internal combustionengine so as to compress the refrigerant, (b) a refrigerant evaporatorfor performing heat exchange with air on the inflowing refrigerant tocause it to evaporate and vaporize, (c) a refrigerant circulationcircuit for circulating the refrigerant discharged by said refrigerantcompressor to said refrigerant evaporator and returning it to saidrefrigerant compressor, (d) variable discharge volume means for reducingthe discharge volume from the refrigeration cycle apparatus when adischarge pressure from said refrigeration cycle apparatus becomeshigher than a predetermined value; and wherein said variable dischargevolume means increases the discharge volume from said refrigerationcycle apparatus when the discharge pressure from said refrigerationcycle apparatus becomes lower than a predetermined value.
 2. Arefrigeration cycle apparatus as set forth in claim 1, furthercomprising a refrigerant passage throttling means provided in the middleof the refrigerant passage directly carrying refrigerant from thedischarge port of the refrigerant compressor to the inlet of saidrefrigerant evaporator for reducing the sectional area of therefrigerant passage when starting up the heating operation where saidrefrigerant circulation circuit is operated.
 3. A refrigeration cycleapparatus comprising: (a) a refrigerant compressor driven in rotation byan internal combustion engine so as to compress the refrigerant, (b) arefrigerant evaporator for performing heat exchange with air on theinflowing refrigerant to cause it to evaporate and vaporize, (c) arefrigerant circulation circuit for circulating the refrigerantdischarged by said refrigerant compressor to said refrigerant evaporatorand returning it to said refrigerant compressor, (d) variable dischargevolume means for increasing the discharge volume from the refrigerationcycle apparatus when a suction pressure into said refrigeration cycleapparatus becomes lower than a predetermined value; and a refrigerantpassage throttling means provided in the middle of the refrigerantpassage directly carrying refrigerant from the discharge port of therefrigerant compressor to the inlet of said refrigerant evaporator forreducing the sectional area of the refrigerant passage when starting upthe heating operation where said refrigerant circulation circuit isoperated.
 4. A refrigeration cycle apparatus as et forth in claim 3,wherein said refrigerant passage throttling means is a valve for openingand closing said refrigerant circulation circuit and said valve closessaid refrigerant circulation circuit after the refrigerant compressor isstarted up until predetermined conditions are satisfied.
 5. Arefrigeration cycle apparatus as set forth in claim 3, wherein saidrefrigerant passage throttling means is a variable throttling valvecomprising an orifice through which the refrigerant passes, a valveelement driving means for reducing the opening degree of the valveelement the lower the discharge pressure of the refrigerant compressor.6. A refrigeration cycle apparatus as set forth in claim 3, wherein saidrefrigerant passage throttling means is a differential pressure valvewhich closes until the discharge pressure of said refrigerant compressorrises over a predetermined value.
 7. A refrigeration cycle apparatuscomprising: (a) a refrigerant compressor driven in rotation by aninternal combustion engine so as to compress the refrigerant, (b) arefrigerant evaporator for performing heat exchange with air on theinflowing refrigerant to cause it to evaporate and vaporize, (c) arefrigerant circulation circuit for circulating the refrigerantdischarged by said refrigerant compressor to said refrigerant evaporatorand returning it to said refrigerant compressor, (d) variable dischargevolume means for increasing the discharge volume from the refrigerationcycle apparatus when a suction pressure into said refrigeration cycleapparatus becomes lower than a predetermined value; and a refrigerantpassage throttling means provided in the middle of the refrigerantpassage directly carrying refrigerant from the discharge port of therefrigerant compressor to the inlet of said refrigerant evaporator forreducing the sectional area of the refrigerant passage when starting upthe heating operation where said refrigerant circulation circuit isoperated; wherein said variable discharge volume means reduces thedischarge volume from said refrigeration cycle apparatus when thesuction pressure into said refrigeration cycle apparatus becomes higherthan a predetermined value.
 8. A refrigeration cycle apparatuscomprising: (a) a refrigerant compressor driven in rotation by aninternal combustion engine so as to compress the refrigerant, (b) arefrigerant evaporator for performing heat exchange with air on theinflowing refrigerant to cause it to evaporate and vaporize, (c) arefrigerant circulation circuit for circulating the refrigerantdischarged by said refrigerant compressor to said refrigerant evaporatorand returning it to said refrigerant compressor, (d) variable dischargevolume means for reducing the discharge volume from the refrigerationcycle apparatus when a discharge pressure from said refrigeration cycleapparatus becomes higher than a predetermined value; and a refrigerantpassage throttling means provided in the middle of the refrigerantpassage directly carrying refrigerant from the discharge port of therefrigerant compressor to the inlet of said refrigerant evaporator forreducing the sectional area of the refrigerant passage when starting upthe heating operating where said refrigerant circulation circuit isoperated.